JP4920708B2 - Positive displacement pump and positive displacement fluid machine equipped with the same - Google Patents

Description

  BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a rolling cylinder type positive displacement pump in which a rolling cylinder rotates together with a revolving revolving piston, and a positive displacement fluid machine including the same, and in particular, a pump operation with a small capacity and low manufacturing cost can be continued smoothly. The present invention relates to a positive displacement pump and a positive displacement fluid machine including the same.

  As a rolling cylinder type positive displacement pump, a configuration in which two or more pistons having different swirl phases are provided in one rolling cylinder as in a hermetic compressor disclosed in Japanese Patent Application Laid-Open No. 11-125191 (Patent Document 1). It was.

Japanese Patent Laid-Open No. 11-125191

  In this rolling cylinder type positive displacement pump, the clearance of each part is 0, and when the turning radius Ep of the rotating piston rotation shaft coincides with the eccentric amount Es of the cylinder rotating shaft, the turning piston swivels at an arbitrary time. The relationship of the following equation (1) always holds between the amount ΔΦp and the rotation amount Φs of the rolling cylinder. In the following description, the amount of rotation of the rolling cylinder will be mainly referred to as the amount of rotation of the stationary body. However, in some cases, the amount of rotation is paraphrased as the amount of rotation. Similarly, with respect to the speed and the central axis, the rotation speed is appropriately referred to as the rotation speed, and the rotation axis is appropriately referred to as the rotation axis.

Φs = Φp / 2 ……………………………………………………………………… (1)
By dividing both sides of the equation (1) by the time t, the equation (1) can be expressed as the following equation (2).

Φs / t = (Φp / t) / 2 (2)
Further, when the turning piston turning speed is expressed as Ωp and the rolling cylinder rotation speed is expressed as Ωs, the expression (1) can be expressed as the following expression (3).

Ωs = Ωp / 2 …………………………………………………………………………………………… (3)
If the relationship of Formula (1) or Formula (3) does not hold for any reason, it will be difficult to recover to a normal relationship. As a result, the pump operation cannot be continued and the displacement pump is locked. End up.

  By the way, in a normal case, the turning motion of the turning piston is realized by the rotation of the crankshaft whose eccentricity is the turning radius. Thus, the rolling cylinder type positive displacement pump has the following characteristics.

  First, since the eccentric pin portion of the crankshaft and the rolling cylinder need to satisfy the relationship of the above formula (1) or formula (3), relative rotation is possible (relative rotation speed is Ωp / 2). The configuration to perform is essential. Specifically, the outer shape of the swivel piston is a cylindrical shape that allows rotation between the outer surface of the swivel piston and the pump groove of the rolling cylinder, or between the inner surface of the swivel piston and the crankshaft eccentric pin part. A configuration that enables rotation is required.

  Secondly, in the rolling cylinder type positive displacement pump, the turning radius Ep of the turning piston and the eccentric amount Es of the rolling cylinder (the amount of eccentricity from the turning center of the turning piston) are the same. Only once, the rotation axis of the orbiting piston coincides with the rotation axis of the rolling cylinder.

  Since the rolling cylinder type positive displacement pump has these characteristics, a situation in which the above formula (1) or (3) is not established in principle is caused by a mechanism as described later. The frequency of occurrence in this case seems to be low, and it has been considered that countermeasures can be taken with a simple response. However, in an actual machine, due to a gap between elements and an error in element arrangement (assembly), a situation in which the above expression (1) or (3) does not hold frequently occurs, and a sampling is performed to avoid locking the pump operation. Experiments have shown that practical measures are essential.

  In the following, the principle mechanism in which the above expression (1) or (3) is not satisfied will be described first, and then the mechanism in the actual machine will be described.

  First, a generation mechanism in which the above formula (1) or (3) is not established in principle will be described with reference to FIG. FIG. 17 shows the positions of the turning piston and the rolling cylinder every 45 degrees when the eccentric pin portion makes one turn clockwise. Here, the outer angle (bold) shown in the figure is the turning amount of the eccentric pin portion, but since it is considered an ideal case where there is no gap between the eccentric pin portion and the turning piston, it is equal to Φp. Become. The inner angle (thin character) is the amount of rotation of the rotating piston's rotation axis (eccentric pin central axis) as seen from the rotation axis of the rolling cylinder. In the ideal case, there is no gap between the rotating piston and the pump groove. Since it is considered, it becomes equal to Φs. At each of these timings, except for the only exception described later, the rotational amount Φs of the rolling cylinder with respect to the rotational amount Φp of the swing piston is determined by the angle shown in FIG. ) Or formula (3) holds.

The exception is the timing (I5 in FIG. 17) at which the cylinder rotation axis (rotation axis) and the orbiting piston rotation axis described above are the second feature of the rolling cylinder type positive displacement pump. Since the rotation axes of the two pump parts coincide with each other, the rolling cylinder does not rotate (Ωp = 0) and the rolling cylinder can rotate (Ωs ≠ 0) from the first feature described above. As a result, the problem (II) can occur through the following situation (I). (I) The pump stops at the timing when the rotation axis of the rolling cylinder coincides with the rotation axis of the orbiting piston.
(II) Some rotation torque is applied to the rolling cylinder, and under the condition (I), only the rolling cylinder rotates without the turning piston turning.

  When the state changes along this mechanism, the situation shown in FIG. 18 is reached where the above formula (1) or (3) is not satisfied. Also, once in this situation, even if the orbiting piston tries to revolve, the force applied to the rolling cylinder is as shown by the arrow in FIG. 19 (any rotation angle of the rotation of the rolling cylinder generated in (II) above). It occurs in the direction passing through the rotation axis of the rolling cylinder. For this reason, torque for turning the rolling cylinder is not generated, and the rolling cylinder cannot rotate and cannot naturally return to the normal posture in which the above formula (1) or formula (3) is established. As a result, the positive displacement pump locks. In order to avoid this pump lock, it has been found that avoiding the above (I) is effective and simple. For example, the set angle of the motor may be determined by using cogging of the motor that is the driving source so that the pump does not stop in the state (I).

  Next, a mechanism for causing the failure of the expression (1) or the expression (3) in the case of an actual machine will be described. According to the above-mentioned principle mechanism, the occurrence of the above (I) is indispensable for the lock, but actually, the lock frequently occurs even if the above (I) does not occur. In other words, there is a lock generation mechanism that cannot be derived from theoretical considerations. For this reason, the mechanism of lock generation in actual machines will be considered again. That is, consider the mechanism that causes the rolling cylinder to deviate from the above formula (1) or (3) during the pump operation.

  The main mechanism caused by the gap between the eccentric pin portion and the turning piston will be described with reference to FIGS. 19 and 20. FIG. 19 is a diagram for explaining the influence of the gap between the eccentric pin portion and the orbiting piston on the pump operation. In FIG. 19, the gap is exaggerated. 19 (R3) to (R5) show the same timing as that of FIGS. 17 (I3) to (I5).

  Since the orbiting piston partitions the suction side pump chamber and the discharge side pump chamber, it receives force from the working fluid in the direction of the arrow in FIG. For this reason, the turning piston is displaced in the direction of the force by the gap between the eccentric pin portion and the turning piston. FIG. 19 shows the displacement in an exaggerated manner, and the rotating piston rotation shaft in that case is indicated by a black circle. Moreover, the eccentric pin part center axis | shaft in each case was also plotted with the rectangle. After FIG. 19 (R5), the eccentric pin portion moves to the lower left side, so the black circles thereafter will always move to the left. As a result, the rotation axis trajectory of the orbiting piston passes through the left side of the rotating cylinder without passing through the rotation axis. FIG. 20 shows an enlarged view of the trajectory of the rotating piston rotation axis.

  Furthermore, as the force acting on the swivel piston, the reaction force of the force applied to the rolling cylinder to rotate the rolling cylinder acts in the direction perpendicular to the pump groove together with the force from the working fluid. Therefore, in FIG. 20, the rotation axis position of the orbiting piston including the displacement due to the reaction force is plotted with white circles. It can be seen that the rotation axis trajectory of the orbiting piston is further shifted to the left by the reaction force from the rolling cylinder. Since the linear direction connecting the rotating piston rotation axis and the cylinder rotation axis is the direction of the pump groove (that is, the rotation phase of the rolling cylinder), the rolling cylinder is in the vicinity (the cylinder rotation axis and the rotation axis of the rotation piston coincide with each other). (Near the position), it can be seen that there is a possibility that a rotation deviating from the above formula (1) or (3) may occur (specifically, the direction of rotation suddenly reverses from clockwise to counterclockwise). . That is, it can be seen that the gap between the eccentric pin portion and the orbiting piston may cause the rotation out of the above formula (1) or formula (3) during the pump operation to lock.

  On the other hand, from this consideration, there is a possibility that even if the rotation axis of the rolling cylinder and the rotation axis of the orbiting piston coincide in principle, even if they do not match, there is a possibility that normal pump operation will occur without locking. It can be explained as follows. That is, if the difference between the rotation axis trajectory of the rolling cylinder and the rotating piston is smaller than the clearance between the pumping groove of the rotating piston and the rolling cylinder, the direction of the pump groove is not changed and the center of the pump groove is deviated. The rotating shaft of the orbiting piston can pass through the position, and the pump operation can be continued.

  The cylinder rotation axis and the rotation axis of the orbiting piston described here mean an instantaneous axis at the moment when the rotation axis of the orbiting piston passes through a position deviated from the center axis of the pump groove.

  As described above, in the case of an actual machine in which there is a gap between pump components, there is a situation in which the above formula (1) or formula (3) does not hold due to the gap that was omitted during the theoretical consideration. At the same time, there is also an effect of correcting the situation in which the expression (1) or (3) is not satisfied due to the gap. The same situation also occurs due to other element gaps not described above and element arrangement errors (for example, the turning radius Ep of the turning piston rotation shaft does not coincide with the eccentric amount Es of the cylinder rotation shaft). In this way, the above-described equation (1) or equation (3) is not satisfied, and the condition of the inter-element clearance and element arrangement where the pump operation lock occurs is determined by a combination of various situations. It is extremely difficult to take countermeasures for each lock occurrence situation.

  Therefore, in order to avoid the operation lock of the rolling cylinder type positive displacement pump and continue the smooth operation, the relationship of the above expression (1) or (3) is always established between the rolling cylinder and the orbiting piston. It is indispensable to provide a rotation regulating means for causing the rotation. At the same time, it is necessary to avoid restricting the relationship of the above formula (1) or (3) too much, and a rotation defining means having the following conditions (A) and (B) is required.

(A) A mechanism other than the current rotating mechanism that defines the rolling cylinder and the revolving piston in the relationship of the above formula (1) or formula (3) is provided.
(B) In the vicinity of the timing when the rotation axis of the rolling cylinder and the rotation axis of the orbiting piston coincide with each other, the regulating action that defines the arrangement of both in the relationship of the above formula (1) or formula (3) is the strongest. As it deviates, its regulatory action is relaxed.

  In Patent Document 1, in order to avoid an operation lock, a countermeasure is taken in which two rolling pistons having a rotation phase different by 180 degrees are provided in one rolling cylinder. This is because, as the rotation axis of one of the revolving pistons approaches the rotation axis of the rolling cylinder, the rotation axis of the other revolving piston is farthest from the rotation axis of the rolling cylinder. The combination serves as a rotation defining means that satisfies the above conditions (A) and (B), which is one of the fundamental measures.

  However, this technique requires two swiveling pistons and is not suitable for reducing the capacity. Furthermore, since the number of parts increases, the processing cost increases. Further, there is a significant decrease in assemblability described below. The crankshaft has two eccentric pin portions eccentric in directions different by 180 degrees, but it is almost impossible to insert the pin into a rolling cylinder in which two pump grooves are set in directions different by 90 degrees. For example, if the eccentric amount and diameter of the eccentric pin portion are designed to be extremely small with respect to the pump groove width, it is possible geometrically, but the outer diameter and height of the swiveling piston must be increased to ensure the amount of seizure. The increase in the load on the eccentric pin portion causes a significant decrease in reliability and performance. For this reason, after dividing a rolling cylinder, after inserting each into the eccentric pin part of a crankshaft, the complicated assembly process of integrating with high position accuracy was needed, and there was a problem of an increase in manufacturing cost. .

  An object of the present invention is to provide a positive displacement pump suitable for a small capacity and capable of reducing manufacturing costs, and a positive displacement fluid machine using the same, while allowing the pump operation to continue smoothly.

A state like the present invention for achieving the above object, a pump casing having a casing chamber, and rolling cylinder which is pivotable arranged in the casing chamber has a pump groove, fitting clearance to the pump groove The two pump chambers are formed by partitioning the pump groove and are arranged so as to be capable of swiveling, and the swiveling speed of the swiveling piston is maintained while maintaining the swiveling of the swiveling piston and the rotation of the rolling cylinder. Rotation regulating means for regulating the rotational speed of the rolling cylinder to twice the rotational speed of the rolling cylinder, the swing piston or the rolling cylinder being rotated or rotated by a drive source, and the rolling cylinder is rotated by the swing piston. With respect to the piston rotation axis that is the central axis of motion, The pump groove is arranged so as to be capable of rotating around an axis, and the pump groove extends linearly across a cylinder rotation axis that is a central axis of the rolling movement of the rolling cylinder. In the positive displacement pump, the swirl radius Ep of the swivel piston and the eccentric amount Es of the rolling cylinder are substantially equal to each other. The number of rolling cylinders is one and the number of swiveling pistons fitted into the rolling cylinder is one, and the rotation defining means synchronizes the rotation of the swiveling piston and the rotation of the rolling cylinder. The rotating piston is provided with a flat side surface that is in sliding contact with the two side surfaces of the pump groove. And means, the piston rotating defining means for defining the rotation speed of the orbiting piston to twice the rotation speed, the piston rotating defining means, on pivoting movement trajectory of the orbiting piston rotation axis, said cylinder rotating A slider mechanism that regulates the movement of the swivel piston so that a stationary point at a position different from the axis is always placed on a straight line that passes through the swivel piston rotation shaft and is fixed on the swivel piston. A rotation guide is provided by a position-fixed cylinder fixedly disposed at a position corresponding to the stationary point of the pump casing; and the rotation guide has the straight line as a center line and the position-fixed cylinder; This is realized by providing a guide groove having a width equivalent to the diameter of the position-fixed cylinder so as to form a sliding pair .

A more preferable specific configuration example in the first aspect of the present invention is as follows.
(1) before SL still point on the pivoting movement trajectory that is provided in a position rotated 180 degrees from the position of the cylinder rotation axis.
( 2 ) The stationary slider includes a fixed central axis and a slider member rotatably inserted into the fixed central axis.
( 3 ) The guide groove is a passage for the working fluid.
( 4 ) The guide groove is extended to the side flat portion.
( 5 ) It is mounted on the positive displacement fluid machine as an oil supply source to each part of the positive displacement fluid machine.
( 6 ) The turning movement of the turning piston is realized by a crankshaft having an eccentricity Ec, the rotating shaft of the rolling cylinder is realized by a bearing portion having an eccentricity Eb, and the rotation piston and the pump groove The gap should be less than twice the difference between the eccentric amount Ec of the crankshaft and the eccentric amount Eb of the bearing portion.

  According to the present invention, it is possible to provide a positive displacement pump suitable for a small capacity and capable of reducing the manufacturing cost, and a positive displacement fluid machine using the same, while allowing the pump operation to continue smoothly.

The longitudinal cross-sectional view of the scroll compressor concerning 1st Embodiment of this invention. The detailed enlarged view of the back pressure chamber vicinity of the scroll compressor of FIG. The expansion longitudinal cross-sectional view of the oil supply pump of the scroll compressor of FIG. KK cross-sectional enlarged view of FIG. VV sectional drawing of FIG. The expanded longitudinal cross-sectional view of the modification 1 near the position fixed cylinder of the oil pump of the scroll compressor of FIG. The expansion longitudinal cross-sectional view of the modification 2 near the position fixed cylinder of the oil pump of the scroll compressor of FIG. The top view of the baseplate of the oil supply pump of the scroll compressor of FIG. The cross-sectional view at the time of changing the installation position of the position fixed cylinder of the oil supply pump of the scroll compressor of FIG. The components expansion | deployment perspective view of the oil supply pump of the scroll compressor of FIG. Operation | movement explanatory drawing of the oil supply pump of the scroll compressor of FIG. Explanatory drawing of the rotation synchronizing means and piston rotation regulation means of the oil pump of the scroll compressor of FIG. The expanded longitudinal cross-sectional view of the position fixed cylinder of the oil pump of the scroll compressor which concerns on 2nd Embodiment of this invention. LL cross-sectional enlarged view of FIG. The expanded cross-sectional view of the position fixed cylinder of the oil supply pump of the scroll compressor which concerns on 3rd Embodiment of this invention. The detailed enlarged view of the back pressure chamber vicinity of the scroll compressor which concerns on 4th Embodiment of this invention. The expanded longitudinal cross-sectional view of the oil supply pump of the scroll compressor which concerns on 4th Embodiment of this invention. Explanatory drawing of the principle pump operation | movement of a rolling cylinder type positive displacement pump. Explanatory drawing of the pump operation which deviated from the principle in a rolling cylinder type positive displacement pump. Explanatory drawing which shows an example of the pump operation | movement which deviated from the principle which actually occurs with a rolling cylinder type positive displacement pump. The figure which shows the turning piston center locus | trajectory at the time of the pump operation | movement which deviated from the principle which actually arises with a rolling cylinder type positive displacement pump. Explanatory drawing of the control action degree of the rolling cylinder rotation angle of the oil pump of the scroll compressor of FIG. The figure which shows the track | orbit of a revolving piston when the clearance gap between the pump groove of the oil pump of a scroll compressor of FIG. 1 and a revolving piston is twice the difference of Eb and Ec.

Hereinafter, a plurality of embodiments in the case where the positive displacement pump of the present invention is mounted as a bearing pump of a scroll compressor which is a positive displacement fluid machine or an oil supply pump to a compression chamber (the working fluid is oil) will be described. It explains using. The same reference numerals in the drawings of the respective embodiments indicate the same or equivalent. In addition, this invention includes making it more effective by combining each embodiment suitably as needed. Further, hereinafter, the working fluid in the positive displacement pump of the present invention is limited to oil, so that it is referred to as oil, and the name of the working fluid refers to the working fluid for the scroll compressor.
(First embodiment)
A first embodiment in which a positive displacement pump according to the present invention is mounted as an oil supply pump in a scroll compressor in which an oil storage part is provided in a casing and the inside of the casing has a suction pressure will be described with reference to FIGS. As a case where such a so-called low pressure chamber type in which the inside of the casing has a suction pressure is adopted, there is a case where a combustible gas is used as a working fluid. For example, hydrocarbon fluids such as propane and butane correspond to this. This is an effective means for reducing the total amount of working fluid enclosed in the entire apparatus including the compressor from the viewpoint of safety.

  First, the overall configuration and operation of the scroll compressor according to the present embodiment, the longitudinal sectional view of the scroll compressor according to the first embodiment of the present invention in FIG. 1, and the vicinity of the back pressure chamber in FIG. 2 (N portion in FIG. 1). This will be described with reference to an enlarged view. In addition, the detail of the oil pump of this embodiment is demonstrated later using FIGS.

  A suction pipe 53 is provided through the side surface of the casing 8, and a working fluid having a suction pressure is introduced into the casing 8 through the suction pipe 53. Then, the working fluid is guided to a compression chamber 100 formed between the fixed scroll 2 and the orbiting scroll 3 through a suction port 2 e opened on the side surface of the fixed scroll 2. Since the compression chamber 100 is reduced in volume while being moved from the outer peripheral portion to the inner peripheral portion by the orbiting motion of the orbiting scroll 3, the working fluid in the compression chamber 100 is compressed. Here, the orbiting motion of the orbiting scroll 3 is realized by rotating the crankshaft 6 connected to the orbiting scroll 3 with the motor 7 and preventing the Oldham ring 5 from rotating.

  The fixed scroll 2 is provided with a bypass valve 22 for avoiding overcompression and liquid compression on the upper surface, and a discharge port 2d through which compressed working fluid is discharged. The fixed scroll 2 is screwed to the frame 4. A back pressure chamber 110 serving as an intermediate pressure (a pressure intermediate between the suction pressure and the discharge pressure, hereinafter referred to as a back pressure) is formed between the back surface of the orbiting scroll 3 and the frame 4.

The crankshaft 6 has an axial position defined by the shaft collar portion 6 h mounted on the frame shaft thrust projection 4 p, and an upper portion is supported by the main bearing 24 and a lower portion is supported by the sub-bearing 25. An eccentric pin portion 6 a provided at the upper end of the crankshaft 6 is inserted into the orbiting bearing 23 of the orbiting scroll 3. Here, the auxiliary bearing 25 includes a ball 25a and a ball holder 25b. The ball holder 25 b is welded to the countershaft support 50 fixed to the casing 8. With the configuration including the ball 25a and the ball holder 25b, the inclination of the countershaft support 50 can be allowed to some extent.

  These bearings are supplied with the oil pumped up by the oil pump 30 from the oil reservoir 125 at the lower part of the casing 8 through the oil hole 6b of the crankshaft 6. The oil supplied to the slewing bearing 23 and the main bearing 24 enters the back pressure chamber 110, and then flows out to the side surface of the frame 4 through the back pressure chamber outflow passage 135 that penetrates the frame 4. Return to 125. Here, a back pressure control valve 26 is provided in the middle of the back pressure chamber outflow passage 135. The back pressure control valve 26 keeps the pressure in the back pressure chamber 110 at a desired back pressure. By this back pressure, the orbiting scroll 3 is urged toward the fixed scroll 2 during compression.

  On the other hand, in order to improve the sealing performance of the compression chamber 100, a small amount of oil is supplied from the slewing bearing chamber 115 to the compression chamber 100 through the compression chamber oil supply passage 130. This oil becomes discharged oil and is discharged to the upper part of the fixed scroll 2 from the discharge port 2d and the bypass valve 22 together with the working fluid.

  A discharge oil separation oil return cylinder 55 is fixed to the upper part of the fixed scroll 2 with a screw to form a discharge chamber 120. A discharge cover 51 having a further protruding discharge pipe 52 is screwed to the upper part of the discharge oil separation oil return cylinder 55 to form an oil separation chamber 90 and an oil return chamber 95.

  The working fluid that has flowed into the discharge chamber 120 is guided to the oil separation chamber 90 to separate oil mixed in the working fluid, and then flows out of the compressor 1 through the discharge pipe 52. The oil separated in the oil separation chamber 90 flows into the oil return chamber 95. The oil that has flowed into the oil return chamber 95 returns to the oil storage portion 125 via the oil return passage 80 connecting the oil return chamber 95 and the oil storage portion 125 and the oil return amount adjustment valve 70 installed in the middle thereof. . The oil return amount adjustment valve 70 keeps a small amount of oil in the oil return chamber 95 in order to seal the discharge side and the suction side on both sides of the oil return passage 80, and the oil amount flowing into the oil return chamber 95. This valve is responsible for returning the same amount to the oil storage part 125. This oil return amount adjustment valve 70 is realized by a float valve or an electromagnetic valve that controls the opening degree by a sensor signal including information on the amount of separated oil.

  The oil supply pump 30 of this embodiment supplies oil in the oil storage part 125 to each bearing part of the crankshaft 6 composed of the auxiliary bearing 25, the main bearing 24, and the slewing bearing 23, and sealability of the compression chamber. In addition to the original role of supplying to the compression chamber 100 for improvement, it also plays a role of supplying to the back pressure chamber 110 to generate back pressure. For this reason, the oil pump 30 is responsible not only for the flow rate but also for increasing the pressure.

  In the present embodiment, a separation oil back pressure chamber introduction passage 500 for introducing a part of the separation oil flowing into the oil return chamber 95 into the back pressure chamber 110 is provided. And the separation oil branching means 501 which adjusts the flow volume which flows oil into this separation oil back pressure chamber introduction channel 500 is provided. Even when the amount of back pressure increase by the oil pump 30 is insufficient, it is possible to increase the back pressure by putting the separated oil of the discharge pressure into the back pressure chamber 110, and avoid the deterioration of the compressor performance due to insufficient back pressure. There is an effect that can be done. As the simplest means for realizing the separated oil branching means 501, there are branch pipes having different pipe diameters.

  The separation oil branching means 501 and the back pressure control valve 26 may be integrated to provide a separation oil introduction back pressure control valve that operates to introduce separation oil into the back pressure chamber 110 when the back pressure does not increase. . In this case, the separated oil is put into the back pressure chamber 110 only when the back pressure is not increased by the oil pump. As a result, there is no need to always put the high-temperature separated oil into the back pressure chamber 110, heating of the compression chamber 100 is suppressed, and the compressor performance is improved.

  Next, the detailed configuration and operation of the oil pump 30 will be described in detail with reference to FIGS. 3 to 11, 21, and 22. The oil supply pump 30 is a rolling cylinder type positive displacement pump having a motor 7 which is a power source of the scroll compressor 1 as a drive source, a revolving piston 30a as a drive side, and a rolling cylinder 30b as a passive side.

  First, the configuration of the oil supply pump 30 will be described with reference to FIGS. 3 to 9. 3 is a longitudinal sectional view of the oil pump 30 (M-part enlarged detail view of FIG. 1, HH sectional view of FIG. 4), and FIG. 4 is a transverse sectional view of the oil pump 30 (KK sectional view of FIG. 3). 5) is a longitudinal sectional view (VV sectional view of FIG. 4) different from FIG. 3 of the oil pump 30, FIG. 6A is an enlarged longitudinal sectional view of a fixing portion between the base plate 30c1 and a separate position fixing cylinder 30p, 6B is an enlarged vertical sectional view of a fixing portion between a base plate 30c1 having a different form from FIG. 6A and a separate position fixing cylinder 30p, FIG. 7 is a plan view of the base plate 30c1, and FIG. 8 is a position and rotation of the position fixing cylinder 30p. 9 is a cross-sectional view of the oil pump 30 when the side surface shape of the piston 30a is changed, and FIG.

  A small-diameter pump shaft portion 6f is provided at the lower end portion of the crankshaft 6, and a pump eccentric portion 6f1 is provided below the pump shaft portion 6f. This pump eccentric portion 6f1 is fitted in a clearance in a piston bearing hole 30a6 formed in the center of the turning piston 30a. The turning piston 30a is swung with a turning radius Ep by the eccentric rotation of the pump eccentric portion 6f1 of the crankshaft 6. An oil supply hole 6b is opened at the lower end surface of the pump eccentric portion 6f1.

  A rotation-free rolling cylinder 30b having a rotation axis γ that is an axis eccentric by an amount of eccentricity Es substantially equal to the turning radius Ep with respect to the turning axis α of the turning piston 30a is provided. The rolling cylinder 30b has a pump groove 30b1 and is disposed in the casing chamber 30c4 of the pump casing 30c so as to be capable of rotating. The pump groove 30b1 extends linearly across the cylinder rotation axis γ, which is the central axis of the rotational motion of the rolling cylinder 30b. The turning piston 30a is fitted in the pump groove 30b1 so as to be reciprocally movable, and the space on both sides of the turning piston 30a is formed as two pump chambers 140 by partitioning the pump groove 30b1. The number of turning pistons 30a fitted to the rolling cylinder 30b is one.

  Here, the distance between the central axis β of the pump eccentric portion 6f1 and the rotation axis α of the crankshaft 6 roughly defines the turning radius Ep of the turning piston 30a. By the way, since there is a gap between the pump eccentric portion 6f1 and the piston bearing hole 30a6, generally a radial force acting on the orbiting piston 30a (with the force described in the problem to be solved by the above-described invention). Yes (see FIGS. 19 and 20), the swivel piston 30a is displaced. Therefore, the interval Ec (notation of Ec) between the center axis β of the pump eccentric portion 6f1 and the crankshaft rotation axis α can be regarded as the center axis of the piston bearing hole 30a6 (this is the rotation axis of the swing piston β And a turning radius Ep that is a distance between the crankshaft rotation axis α and the crankshaft rotation axis α.

As is apparent from FIG. 9, the orbiting piston 30 a has a flat end 30 a 5 at one end face (lower end face), and connects two parallel side face flat parts 30 h and these two side face flat parts 30 h. Two side cylindrical surfaces 30a4 are provided on the side surfaces. Two side cylindrical surface 30a4, as is apparent from FIG. 8, by shifting the mutual axis, are made to coincide with the cylindrical surface forming a cylinder chamber 30C4. Thereby, since the volume of the pump chamber 140 formed is reduced to 0 in principle, there is an effect that the dead volume is eliminated and the performance is improved.

  The planar end portion 30a5 has a guide groove 30g that extends between the two side flat portions 30h and allows the lower surface lower space and the piston bearing hole 30a6 to communicate with each other. The width of the guide groove 30g is constant. As described above, the rotation-free rolling cylinder 30b is provided in which the rotation axis γ is an axis that is eccentric from the piston rotation axis α by approximately the same eccentric amount Es as the turning radius Ep of the turning piston 30a. As a result, as shown in FIG. 8, the cylinder rotation axis γ comes on the turning locus of the turning piston rotation axis β ′. As a result, the timing at which the rotation piston β ′ and the cylinder rotation axis γ coincide with each other only once while the rotation piston 30a makes one rotation.

  The rolling cylinder 30b has an end plate portion 30b4 at the upper part in the axial direction and a pump groove 30b1 at the lower part in the axial direction. The swing piston 30a is mounted in the pump groove 30b1 so that the side flat portion 30h of the swing piston 30a and the side surface of the pump groove 30b1 are in sliding contact and reciprocate. Thereby, the pump groove 30b1 is partitioned into two spaces, and each space becomes the pump chamber 140. The pump chamber 140 is separated from the internal space of the casing 8 by the pump casing 30c.

  The pump casing 30c includes a base plate 30c1 and a cover 30c2 provided on the lower surface and the upper surface side of the rolling cylinder 30b, respectively, and a pump cylinder 30c3 which is a connecting portion thereof and which rotatably supports the rolling cylinder. As shown in FIG. 7, the base plate 30c1 has a suction dig 30s1 provided on the upper surface and a pump suction hole 30s2 penetrating the base plate 30c1, thereby providing a pump suction passage 30s for sucking oil from the oil storage part 125. It is composed. Further, the base plate 30c1 has a pump discharge dig 30d1 and a pump discharge groove 30d2 provided on the upper surface, and thereby constitutes a pump discharge flow path 30d for sending oil to the oil supply hole 6b opened at the lower end of the crankshaft 6. is doing.

A position fixing cylinder 30p protruding upward is provided integrally with the base plate 30c1 at the center of the base plate 30c1. The fixed position cylinder 30p is provided at a position rotated 180 ° from the cylinder rotation axis γ on the turning locus of the rotation axis β ′ of the turning piston 30a. Incidentally, as shown in FIGS. 6A and 6B, stationary cylinder 30p separately from the base plate 30c1 0 -1,30p 0 -2 may be provided. The position fixing cylinder 30p 0 -1 is formed of a cylinder having the same diameter, and is press-fitted into the hole of the base plate 30c1 from above and fixed. The position fixing cylinder 30p 0 -1 is formed of a cylinder having a flange at the lower end, and is press-fitted from below into the hole of the base plate 30c1 and fixed.

  In the pump casing 30c of this embodiment, as can be seen from FIGS. 3 and 5, an upper pump casing 30c23 in which a cover 30c2 and a pump cylinder 30c3 are integrated is used. As a result, the number of parts is reduced, and the assemblability can be improved. The upper pump casing 30c23 is provided with a minimum necessary hole for passing the pump eccentric portion 6f1.

  As shown in FIG. 9, the oil pump 30 is actually assembled by first passing the pump eccentric portion 6f1 through the hole of the upper pump casing member 30c23, and then incorporating the rolling cylinder 30b and the turning piston 30a. Temporarily fixed to the holder 25b (which may be the auxiliary bearing support plate 50).

In this state, the base plate 30c1 is fixed to the upper pump casing 30c23 via the base plate fixing screw 30m while the position fixing cylinder 30p is inserted into the guide groove 30g. At this time, by inserting a knock pin as a positioning hole 30i 1 positioning holes 30i2 and the upper pump casing 30c23 of the base plate 30c1 fit increases the set position accuracy of the pump suction passage 30s and the pump discharge passage 30d.

  Thereafter, while rotating the crankshaft 6, the pump cylinder fixing screw 30k is finally tightened to fix the pump cylinder 30c3 to the ball holder 25b. Thereby, since the positional accuracy of the pump cylinder 30c3 can be increased, the positional accuracy of the cylinder rotation shaft γ is improved, the operation of the oil pump 30 can be smoothed, and the performance of the oil pump is improved. Here, as a method of rotating the crankshaft 6, the motor 7 can be rotated at a low speed, or the orbiting scroll member 3 can be revolved by sucking air from the suction port 2e with a vacuum pump.

  Next, operation | movement of the oil supply pump 30 is demonstrated using FIG.10, FIG.11, FIG.21 and FIG. FIG. 10 is a diagram for explaining the operation of the oil supply pump 30, and shows the pump operation in the same cross section as in FIG. 4 while the pump chamber 140 proceeds in one stroke. FIG. 11 is a diagram for explaining the operation when the rotation centers of the swing piston 30a and the rolling cylinder 30b coincide with each other, FIG. 21 is a diagram for explaining the degree of regulation of the rolling cylinder rotation angle, and FIG. 22 is the gap between the pump groove 30b1 and the swing piston 30a. Is the trajectory of the orbiting piston 30a when there is twice the difference between Eb and Ec.

  As shown in FIG. 10, during one stroke of the pump chamber 140, the crankshaft 6 rotates twice (in the direction of a circular arrow). Note that FIG. 10 shows a cross-sectional change every time the crankshaft 6 rotates 22.5 degrees, and hatching that represents a cross section of each component is omitted.

  As described above, two pump chambers 140 are formed at the same time. Since these two pump chambers 140 are out of phase with each other, when one pump chamber is in the suction stroke, the other pump chamber is in the discharge stroke, but the operation changes are the same. Therefore, paying attention to one pump chamber (cross-hatched pump chamber in FIG. 10), the pump operation will be described. When the pump chamber 140 is in the suction stroke, it is called a suction pump chamber 140s, and when it is in the discharge stroke, it is called a discharge pump chamber 140d.

  The suction strokes (1) to (10) in FIG. 10 (numbers in the drawing are indicated by parentheses in the specification) are the suction strokes, and the oil in the oil storage section 125 passes through the suction passage 30s and the suction pump chamber 140s. It is a process of sucking up. 10 (11) to (16) is a discharge stroke, and is a step of discharging the oil in the discharge pump chamber 140d to the oil supply hole 6b through the discharge flow path 30d and the guide groove 30g. The pump operation of the conventional configuration (the configuration without the fixed position cylinder 30p, the guide groove 30g and the side flat portion 30h newly installed in the present embodiment) (when the gap between each portion is extremely small and there is no assembly error) and The problems have already been explained in the problem to be solved by the invention (see FIGS. 17 and 18), and further, the actual pump operation and problems with gaps in each part and errors in assembly are also explained in FIGS. It is done.

  Therefore, the following description will be focused on the position-fixed cylinder 30p, the guide groove 30g, and the side flat portion 30h that are newly installed in the present embodiment, using FIG. 10 and FIG. That is, it will be described that the position fixing cylinder 30p, the guide groove 30g, and the side flat portion 30h serve as rotation defining means that satisfies the conditions (A) and (B). First, it will be explained that a mechanism that satisfies the conditions (A) and (B) can be configured by the elements newly installed in the present embodiment, and then the conditions (B) are obtained by optimizing the installation locations of these elements. Will be better satisfied.

  In order to satisfy the condition (A) “always regulate the turning speed of the turning piston 30a to be twice the rotation speed of the rolling cylinder 30b”, first, the rotation that synchronizes the rotation speed of the rolling cylinder 30b with the rotation speed of the turning piston 30a. By providing the synchronization means, the condition (A) is converted into a prescribed condition for only the turning piston 30a. Specifically, it is changed to the content of providing a piston rotation defining means for defining the turning speed of the turning piston 30a to be twice the rotation speed.

  The rotation synchronizing means of the present embodiment is realized by providing two side flat portions 30h on the side surface of the turning piston 30a and inserting the side piston 30a so as to be in sliding contact with the pump groove 30b1 of the rolling cylinder 30b. This rotation synchronization means is a seal portion that partitions the suction pump chamber 140s and the discharge pump chamber 140d, and has a surface seal instead of a conventional line seal, and thus has an effect of improving the sealing performance.

  The piston rotation defining means, which is another constituent means, is a guide groove (guide groove installation angle) through which the position-fixed cylinder 30p disposed on the rotation path of the rotation axis β ′ of the rotation piston 30a passes through the rotation piston rotation axis β ′. Is realized by a slider mechanism that is inserted into 30 g. The fact that this slider mechanism becomes a piston rotation defining means means that the oil supply pump can be operated as shown in FIG. 10 even if a rotation defining means configured by combining the slider mechanism and the synchronous rotation means is set. It is clear from this.

  In this way, the mechanism combining the side flat portion 30h that realizes the rotation synchronizing means and the slider mechanism that is constituted by the position fixing cylinder 30p and the guide groove 30g that realizes the piston rotation defining means satisfies the condition (A). Fulfill.

  Next, it will be described that the condition (B) is satisfied, but before that, the prescribed working degree of the rotation regulating means is defined. The prescribed degree of action is an index of how accurately it can be defined with respect to the target definition, and is defined as the reciprocal of the difference between the target value and the actual set value. The target value of the present embodiment is indicated by the above formula (1), and the reciprocal of the error (rotation) between the target rotation amount of the rolling cylinder and the actual rotation amount corresponding to the turning amount Φp of the turning piston is the prescribed action. Degree. That is, the following formula (4) is obtained.

Specified degree of action ≡1 / rotation …………………………………………………………………………………… (4)
The main cause of the rolling cylinder rotation amount error (rotation) is considered to be the clearance (back) between the fixed position cylinder 30p and the guide groove 30g. Therefore, an equation for the prescribed working degree is obtained by setting this rotation to Δ. As shown in FIG. 21, the parameters of the prescribed working degree include a set angle of the position-fixed cylinder 30p (the angle on the trajectory of the turning piston rotation axis β ′ is θ) and a turning amount of the turning piston 30a (Φp, The definition is the same as in FIG. In FIG. 21, for the sake of explanation, the rotation angle Δ is enlarged, and the diameter of the fixed position cylinder 30p is reduced. From FIG. 21, the rotation Δ is obtained from the circumferential deviation Δs at the radius L as the following equation (5).

Rotation Δ ＝ Arctan (abs (Δs / L)) ………………………………………………………………………………………………………………………………………………………………………… (5)
Here, abs represents an absolute value.

L and Δs in the equation (5) are the following equations (6) and (7).
L = 2 ・ Es ・ cos (θ / 2) ………………………………………………………………………………………… (6)
Δs = Δ / cos (χ) …………………………………………………………………………………………………………………… (7)
Here, χ is also an angle formed by a guide groove 30g and a line connecting the cylinder rotation axis γ and the position fixing cylinder 30p as shown in FIG.

The guide groove 30g passes through the cylinder rotation shaft γ when the orbiting piston rotation axis β ′ reaches the position of the cylinder rotation axis γ, the rotation speed of the guide groove 30g is synchronized with the rolling cylinder 30b, and the rolling cylinder 30b Since the rotational speed is half of the rotational speed of the swing piston 30a, it can be seen that χ is given by the following equation (8).
χ = abs (π−Φp) / 2 ………………………………………………………………………………………………………… (8)
Substituting equation (8) into equation (7), and then substituting into equation (5) together with equation (6), finding the rotation Δ, and finally substituting into equation (4), If it calculates | requires, it will become the following formula | equation (9).
Specified degree of action = 1 / (Arctan (Δ / (2Ep · abs (cos (θ / 2) · sin (Φp / 2)))))) (9)
When the prescribed working degree at Φp = π rad in which the rotating piston rotation axis β ′ coincides with the cylinder rotation axis γ and Φp = 0 rad rotated by π rad (180 degrees) than that is calculated by the equation (9), the following equation (9) 10) and (11).
Normal degree of action (Φp = π) = 1 / (Arctan (Δ / (2Ep · abs (cos (θ / 2))))) (10)
Normal degree of action (Φp = 0) = 1 / (Arctan (∞)) = 2 / π (11)
Except for the case of θ = πred, the following inequality (12) holds.
Normality (Φp = π)> Normality (Φp = 0) (12)
As a result, the prescribed working degree of the newly set rotation regulating means is low (the turning piston) by the driving system except when the fixed position cylinder 30p coincides with the cylinder rotation axis γ (θ = πrad). High when the rotation axis β ′ coincides with the cylinder rotation axis γ), and low when the specified operation by the drive system is high (the rotation piston rotation axis β ′ turns 180 degrees from the position where it coincides with the cylinder rotation axis γ). I understand that. From this, it can be seen that the condition (B) is satisfied.

  From the above, even if the position-fixed cylinder 30p is set at any position different from the cylinder rotation axis γ, the newly provided piston-to-cylinder rotation restricting means cooperates with the drive system and has a low prescribed working degree. In contrast, when the specified working degree is high, the prescribed working degree is lowered so as not to be over-constrained, and an effect of realizing a smooth pump operation is achieved.

  Further, as is clear from the equation (9), as the set angle θ is changed from π rad to 0 rad, the absolute value of the specified working degree increases (when the turning amount Φp is the same), and it becomes maximum at 0 rad. That is, when the fixed position cylinder 30p is provided at a position rotated 180 degrees on the trajectory of the orbiting piston rotation axis β ′ from the position of the cylinder rotation axis γ, the specified working degree becomes the highest and the rotation Δ can be suppressed. The pump operation can be made smoother.

  The position fixing cylinder 30p is erected in the pump discharge passage 30d as is apparent from FIGS. Further, the guide groove 30g serves as a discharge path connecting the pump discharge flow path 30d and the oil supply hole 6b. Therefore, the sliding portion between the fixed position cylinder 30p and the guide groove 30g is always located at the center of the oil passage, so that the lubricity is improved and the reliability is improved.

  In this embodiment, the guide groove 30g extends to the side flat portion 30h. As described above, since the guide groove 30g is an oil passage, the oil is present in the guide groove 30g. Therefore, the guide groove 30g extending to the side flat portion 30h serves as an oil supply passage to the side flat portion 30h, and has an improved reliability due to improved lubricity between the side flat portion 30h and the pump groove 30g, and has a sealing property. With the improvement, the performance improvement effect is achieved by reducing the leakage from the discharge pump chamber 140d to the suction pump chamber 140s. Further, by extending the guide groove 30g to the side flat portion 30h, the movement of the cutting tool at the time of grooving becomes uniform, and the shape accuracy of the groove is improved.

  A conforming film such as manganese phosphate may be provided on the surface of the revolving piston 30a, the rolling cylinder 30b, or the pump casing 30c. In this case, due to the familiar effect, each gap is reduced with operation, and the sealing performance is improved and the pump performance is improved.

  Further, in this embodiment, the swing piston 30a is the drive side and the rolling cylinder 30b is the passive side, but conversely, the rolling cylinder 30b may be the drive side and the swing piston 30a may be the passive side. In this case, the turning piston 30a performs crank support so that the turning motion can be performed.

  In the present embodiment, the pump groove 30b1 extends to the outer periphery of the rolling cylinder 30b. Thereby, since the movement of the cutting tool during groove processing becomes uniform, there is an effect that the shape accuracy of the groove is improved. In addition, since oil is supplied to the outer peripheral clearance of the rolling cylinder, the sealing performance of the outer peripheral clearance can be improved, so that the pump performance is improved. Further, the end plate 30b4 is required for the penetrating pump groove 30b1, and thereby the configuration of a fourth embodiment described later that can suppress the side gap by pressing is configured. Of course, the pump groove type of the conventional long hole shape without penetrating to the outer periphery may be used.

  Moreover, in this embodiment, although the axial center of the two cylindrical surfaces 30a4 provided on the side surface of the orbiting piston 30a is shifted, as the same cylindrical surface shown by the one-dot chain line in FIG. Also good. The reason for this is that the working fluid of the present embodiment is an incompressible fluid called oil, so there is little performance degradation due to dead volume, and the side surface shape is reduced by cutting the side flat portion 30h after forming a cylindrical surface. Since it can process, there exists an effect that processing cost reduces. In addition, in the case of an application in which only oil transfer is performed without increasing the pressure, the performance degradation due to the dead volume is further reduced to a negligible level. Therefore, the swiveling piston 30a having the same cylindrical surface shown by the one-dot chain line in FIG. 8 is suitable. ing.

  In this embodiment, the fixed position of the position fixing cylinder 30p is a slide mechanism provided at a position rotated 180 degrees from the position of the cylinder rotation axis γ on the turning locus of the turning piston rotation axis β ′. As described in the explanation of the degree, only half the movement angle of the moving position fixed cylinder 30p ′ moved to the other position on the turning locus of the rotating piston rotation axis β ′ and the moving position fixed cylinder 30p ′ is set. Even if the moving guide groove 30g ′ whose angle is rotated is provided, the slider mechanism serving as the piston rotation defining means can be configured (see FIG. 10). For example, as shown in FIG. 8, when the position fixing cylinder 30p is set at a position rotated by 45 degrees, the guide groove rotates by 22.5 degrees. In this case, as can be seen from the above description of the degree of restriction action, the timing at which the degree of restriction action becomes maximum is the same as that before the movement of the position-fixed cylinder ((5) in FIG. 10). However, as can be seen from the equation (9), the degree of regulation is smaller than that before the fixed position column 30p is moved. Accordingly, when it is desired to reduce the degree of regulation for reasons such as reducing the load on the sliding portion, or when the position fixing cylinder 30p is positioned on the turning locus of the rotating piston rotation axis β ′ due to restrictions on the arrangement of parts, the position of the cylinder rotation axis γ If it cannot be installed at a position rotated 180 degrees from the movement position, the movement position fixed cylinder 30p ′ and the movement guide groove 30g ′ may be provided.

  In the embodiment described so far, the turning radius Ep of the turning piston 30a and the rolling cylinder are set in the vicinity of the rotation axis β ′ of the turning piston that is likely to cause abnormal rotation of the rolling cylinder 30b that induces the lock and the cylinder rotation axis γ. It was assumed that the eccentric amount Es of 30b is substantially equal (when the rotating piston rotation axis β ′ passes through the cylinder rotation axis γ). That is, the case where the turning radius Ep and the eccentricity Es are different has not been considered. This is due to the following reason.

  The turning radius Ep is determined by the amount of eccentricity Ec of the pump eccentric portion 6f1 and the amount of eccentricity of the bearing clearance in the piston bearing hole 30a6. It is determined by the amount of eccentricity of the outer peripheral bearing clearance of 30c3. The rotation defining means of the present embodiment is means for forcibly changing the amount of bearing eccentricity in both directions so that the trajectory of the rotating piston rotation axis β ′ near the cylinder rotation axis γ passes through the cylinder rotation axis γ. In other words. In other words, as a result of the provision of the rotation defining means, only when the turning radius Ep is equal to the eccentricity Es in the vicinity where the turning piston rotation axis β ′ and the cylinder rotation axis γ are closest to each other.

  The change in the amount of eccentricity of the bearing by the rotation defining means described in the previous paragraph becomes larger when the difference between the amount of eccentricity Ec and the amount of eccentricity Eb is large, and the load applied to the defining means and the pump eccentric portion 6f1 increases. Reliability decreases. For this reason, in the present embodiment, the gap between the orbiting piston 30a and the pump groove 30g is set to twice the difference between the eccentric amount Ec and the eccentric amount Eb. As a result, even if the turning radius Ep and the eccentric amount Es are different in the vicinity where the turning piston rotation axis β ′ and the cylinder rotation axis γ are closest, as shown in FIG. 22, the turning piston 30a remains in the center in the pump groove 30b1. Although deviating from the axis, the rolling cylinder 30b is not abnormally rotated. As a result, it is not necessary to change the amount of bearing eccentricity described above, and the load applied to the defining means and the pump eccentric portion 6f1 does not increase, and there is an effect that reliability can be ensured.

  When the clearance between the swing piston 30a and the pump groove 30g is set to be twice or more the difference between the eccentric amount Ec and the eccentric amount Eb, the clearance between the swing piston 30a and the pump groove 30b1 is enlarged, so that the sealing performance is reduced. Performance is reduced. From this, it can be seen that the clearance between the turning piston 30a and the pump groove 30g may be made equal to or less than twice the difference between the eccentric amount Ec and the eccentric amount Eb.

  According to the present embodiment, since the pump chamber with high sealing performance can be configured while keeping the processing cost low with a simple element shape, there is a risk of operation lock in a low-cost and high-performance rolling cylinder type positive displacement pump. Can be avoided with the configuration of one piston per cylinder. As a result, a positive displacement pump suitable for a small capacity can be realized. Further, compared to the conventional multiple cylinders, the configuration is simple and the assemblability is improved, so that a positive displacement pump with reduced machining cost can be realized.

(Second Embodiment)
Next, about the scroll compressor 1 which mounts the positive displacement pump which is the 2nd Embodiment of this invention as the oil supply pump 30, FIG. 12 which is an expansion longitudinal cross-sectional view of the position fixed cylinder 30p, and its cross section in LL This will be described with reference to FIG. The second embodiment is different from the first embodiment in the points described below, and the other points are basically the same as those in the first embodiment, and thus redundant description is omitted.

  In the second embodiment, a position fixing column 30p is composed of a center pin 30P3 fixed to the base plate 30c1 and a cylindrical roller 30p4 by press fitting, adhesion, or electrodeposition. A flange portion 30p31 is provided at the upper end of the center pin 30p3. The roller 30p4 is fitted to the main body protruding portion 30p32 of the center pin 30P3, and is prevented from being removed from the center pin 30p3 by the flange portion 30p31. In FIG. 12 and FIG. 13, the radial gap between the roller 30p4 and the center pin 30p3 is drawn large, but in actuality it is 100 μm or less. Accordingly, since the roller 30p4 rotates so that the relative speed of the portion where the roller 30p4 and the guide groove 30g slide strongly becomes smaller, the sliding state between the position fixing column 30p and the guide groove 30g becomes good, and chatter occurs. Such an unsatisfactory movement is suppressed, and an effect of smoothing the pump operation of the oil supply pump 30 is achieved. Further, wear of the position fixing cylinder 30p and the guide groove 30g, dropout of the position fixing cylinder 30p due to the twisting and the like can be avoided, and the reliability of the oil supply pump 30 is improved.

(Third embodiment)
Next, the scroll compressor 1 in which the positive displacement pump according to the third embodiment of the present invention is mounted as the oil supply pump 30 will be described with reference to FIG.

  The third embodiment is the same as the second embodiment except that the slider roller 30p5 is provided with a roller flat surface 30p51 on the outer peripheral surface, and the other description is omitted. According to this 3rd Embodiment, since the sliding part area with the guide groove 30g increases, there exists an effect that the danger of abrasion falls and the highly reliable oil supply pump 30 can be provided.

(Fourth embodiment)
Next, the scroll compressor 1 which mounts the positive displacement pump which is the 4th Embodiment of this invention as the oil supply pump 30 is demonstrated using FIG. 15, FIG.

  In the fourth embodiment, the axial position of the crankshaft 6 is received on the oil pump 30 side instead of the frame 4, and the first to third embodiments except for the vicinity of the back pressure chamber 110 and the oil pump 30 are used. It is the same as the form. Therefore, mainly using FIG. 15, which is a detailed enlarged view of the vicinity of the back pressure chamber (corresponding to the N portion in FIG. 1), and FIG. 16, which is an enlarged vertical sectional view of the oil pump (corresponding to the M portion in FIG. 1). It will be described, and other descriptions will be omitted.

In the fourth present embodiment, (see FIG. 15) System Yafuto flange portion 6h is separated from the frame 4, the first through third oil supply pump 30 which was in the embodiment of the top cover (30c2 in FIG. 5) is missing (See FIG. 16) A shaft lower end surface 6z that is an end portion of the crankshaft 6 is supported by an end plate 30b4 of the rolling cylinder 30b. The crankshaft 6 has an upper region facing the back pressure chamber 110, while a lower portion has a region facing the suction pressure, which is the internal space pressure of the casing 8, so that a force that pushes down is always applied. Therefore, the crankshaft 6 presses the end plate 30b4 downward. Therefore, the gap between the upper surface of the turning piston 30a and the bottom surface of the pump groove 30b1, which is the side gap of the oil pump 30, the gap between the flat end 30a5 (the lower surface of the turning piston) and the upper surface of the base plate 30c1, and the upper surfaces of the rolling cylinder 30b and the base plate 30c1. The gap can be reduced. Among these, the smallest gap is almost zero and slides. Which gap is slid is determined by the magnitude relationship between the depth of the pump groove 30b1 and the thickness of the orbiting piston 30a.

  When the thickness of the turning piston 30a is made smaller than the depth of the pump groove 30b1, the gap between the rolling cylinder 30b and the upper surface of the base plate 30c1 slides. In this case, since the revolving piston 30a that performs the revolving motion becomes smooth, there is an effect that the pump operation of the entire oil supply pump 30 becomes even smoother. On the contrary, when the thickness of the swing piston 30a is larger than the depth of the pump groove 30b1, the gap between the upper surface of the swing piston 30a and the bottom surface of the pump groove 30b1, the planar end 30a5 (the lower surface of the swing piston) and the upper surface of the base plate 30c1. The gap becomes a sliding surface. In this case, since the gaps at the two locations can be suppressed to the minimum value, the sealing performance can be further improved, so that an effect of providing a high-performance oil pump is obtained.

  Further, as shown in FIG. 16, the entire oil pump 30 is fixed to the lower surface of the ball 25 a constituting the auxiliary bearing 25 by a pump fixing screw 30 n. Here, the lower surface of the ball 25a has a squareness with respect to the inner peripheral surface which is a bearing surface. Further, it is assumed that the shaft lower end surface 6z is perpendicular to the center axis of the crankshaft 6. Thereby, regardless of the mounting posture of the crankshaft 6, the shaft lower end surface 6z is in contact with the end plate 30b4 over the entire surface, so that the sealing performance between the shaft lower end surface 6z and the end plate 30b4 is improved, and a high-performance oil pump There is an effect of providing. Moreover, since the one-side contact between the shaft lower end surface 6z and the end plate 30b4 is suppressed, there is an effect that a highly reliable oil supply pump can be provided.

  DESCRIPTION OF SYMBOLS 1 ... Scroll compressor, 2 ... Fixed scroll member, 3 ... Orbiting scroll member, 4 ... Frame, 5 ... Oldham ring, 6 ... Crankshaft, 6b ... Oil supply hole, 6f ... Pump shaft part, 6f1 ... Pump eccentric part, 7 ... Motor, 22 ... Bypass valve, 26 ... Back pressure control valve, 30 ... Oil pump, 30a ... Revolving piston, 30a5 ... Planar end, 30a6 ... Piston bearing hole, 30b ... Rolling cylinder, 30b1 ... Pump groove, 30b4 ... End Plate, 30c ... Pump casing, 30c4 ... Casing chamber, 30d ... Pump discharge flow path, 30g ... Guide groove, 30h ... Side flat part, 30p ... Position fixed cylinder, 30p4 ... Roller, 30p5 ... Slider roller, 30s ... Pump suction flow 100, compression chamber, 105 ... suction chamber, 110 ... back pressure chamber, 120 ... discharge chamber, 125 ... oil storage section, 14 ... Pump chamber, 140s ... Suction pump chamber, 140d ... Discharge pump chamber, α ... Piston rotation axis (crankshaft rotation axis), β ... Pump eccentric part central axis, β '... Rotation piston rotation axis, γ ... Cylinder rotation axis ( Cylinder rotation axis).

Claims (7)

A pump casing having a casing chamber;
A rolling cylinder having a pump groove and arranged in the casing chamber for rotational movement;
A swiveling piston that is fitted in the pump groove so as to form two pump chambers by partitioning the pump groove and is arranged to be capable of swiveling,
Rotation regulating means for maintaining the rotation of the swiveling piston and the rotation of the rolling cylinder and defining the swirling speed of the swirling piston to be twice the rotation speed of the rolling cylinder,
The swiveling piston or the rolling cylinder is swung or rotated by a driving source,
The rolling cylinder is disposed so as to be capable of rotational movement about a cylinder rotational axis having an eccentricity of Es with respect to a piston rotational axis that is a central axis of the rotational movement of the rotational piston.
The pump groove extends linearly across the cylinder rotation axis, which is the central axis of the rotational movement of the rolling cylinder,
The revolving piston reciprocates in the pump groove, and is arranged so as to be capable of rotating with a revolving motion with a revolving radius of Ep with respect to the pump casing.
In the positive displacement pump, the turning radius Ep of the turning piston and the eccentric amount Es of the rolling cylinder are substantially equal.
The number of the rolling cylinders is one and the number of the swiveling pistons fitted to the rolling cylinder is one,
The rotation regulating means is a rotation synchronizing means provided with side flat portions that are in sliding contact with the two side surfaces of the pump groove on the turning piston so as to synchronize the rotation of the turning piston and the rotation of the rolling cylinder. And a piston rotation defining means for defining the turning speed of the turning piston to be twice the rotation speed ,
The piston rotation defining means is always placed on a straight line that is fixed on the swivel piston passing through the swivel piston rotation axis on a trajectory of movement of the swivel piston rotation shaft and having a stationary point different from the cylinder rotation axis. A slider mechanism for regulating the movement of the swivel piston,
The slider mechanism includes a stationary slider and a rotation guide,
The stationary slider is realized by a position-fixed cylinder fixedly disposed at a position corresponding to the stationary point of the pump casing,
The rotation guide is realized by providing a guide groove having a width equal to the diameter of the position fixing column to form a sliding pair with the position fixing column with the straight line as a center line on the orbiting piston ,
The positive displacement pump. In claim 1,
Providing the stationary point at a position rotated 180 degrees from the position of the cylinder rotation axis on the trajectory of movement;
The positive displacement pump. In claim 2,
The stationary slider comprises a fixed central axis and a slider member rotatably inserted in the fixed central axis;
The positive displacement pump. In claim 3,
The guide groove as a working fluid passage;
The positive displacement pump. In claim 4,
Extending the guide groove to the flat side surface;
The positive displacement pump. In any of claims 1 to 5,
It is mounted on the positive displacement fluid machine as an oil supply source to each part of the positive displacement fluid machine,
The positive displacement pump. In any one of Claims 1 thru | or 6,
The swing movement of the swing piston is realized by a crankshaft whose eccentricity is Ec,
The rotating shaft of the rolling cylinder is realized with a bearing portion whose eccentricity is Eb,
The clearance between the orbiting piston and the pump groove is not more than twice the difference between the eccentric amount Ec of the crankshaft and the eccentric amount Eb of the bearing portion,
The positive displacement pump.

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